Engine apparatus

ABSTRACT

The fuel-per-stroke which fuel injection systems deliver to an engine is varied in accordance with engine speed to provide a desired fuel-air mixture ratio by means of a spring-loaded check valve responsive to momentary pressure impulses occurring in a pump chamber and operative to divert increasing fuel at increasing engine speeds. The check valve spring loading is varied simultaneously with the pump delivery setting in some embodiments, and valve operation is varied in accordance with engine speed and/or acceleration in some embodiments. The system is illustrated in connection with several reciprocating fuelinjection pumps and several rotary fuel-injection pumps.

[ June 19, 1973 ENGINE APPARATUS Otmar M. Ulbing, R.D. No. l, Berkshire,NY.

[22] Filed: June 21, 1971 [21] Appl. No.: 155,114

Related US. Application Data [76] Inventor:

[52] US. Cl. 137/565.2, 417/255, 417/289,

[51 int. C1. F04b 49/00, F04b 39/00 [58] Field of Search 417/289, 303,310,

[56] References Cited UNITED STATES PATENTS 2,117,512 5/1938 Scott417/462 3,057,300 10/1962 Ulbing 137/565.l

3,120,814 2/1964 Mueller 417/310 Primary Examiner-William L. FreehAttorney-Richard G. Stephens [5 7] ABSTRACT The fuel-per-stroke whichfuel injection systems deliver to an engine is varied in accordance withengine speed to provide a desired fuel-air mixture ratio by means of aspring-loaded check valve responsive to momentary pressure impulsesoccurring in a pump chamber and operative to divert increasing fuel atincreasing engine speeds. The check valve spring loading is variedsimultaneously with the pump delivery setting in some embodiments, andvalve operation is varied in accordance with engine speed and/oracceleration in some embodiments. The system is illustrated inconnection with several reciprocating fuel-injection pumps and severalrotary fuel-injection pumps.

28 Claims, 19 Drawing Figures Patented June 19, 1973 10 Sheets-Sheet 1ni F INVENTOR.

I OTMARH-ULBING Way Patented June 19, 1973 10 Sheets-Sheet 2 PatentedJune 19, 1973 3,739,309

10 Sheets-Sheet 3 I MINIMUM DELIVERY III III MEDIUM MAXIMUM DELIVERYDELIVERY Patented June 19, 1973 10 Sheets-Sheet 4 NQK Patented June 19,1973 3,739,809

10 Sheets-Sheet 5 Patented June 19, 1973 10 Sheets-Sheet 6 In 2 E FPatented June 19, 1973 3,739,809

10 Sheets-Sheet '7 a! c I d e f K r1 :1 FL FL FL 11 I I g h I I k J J TL1 FL FL m I z 1 :12; b m c n: d I::! e :3 f I H I I 9 h J I: k

I o. b c d f G l l l I l l g h i k F -1 l 'lT [T m W l E 3 7 ha 2 a: D'Z} q h E i C: j 2:: k 4:: ffl L1 I I J! L J I JTL -J W REV I PLATE225:! 1:: I: L71: II': I: A 7- I *r- I W 9 0 IB O 2 /0 36 0 PatentedJune 19, 1973 10 Sheets-Sheet 8 ENGINE FIG. 30.

CO AP NOZZLE SUPPLY Patented June 19, 1973 3,739,809

10 Sheets-Sheet 9 Patented June 19, 1973 10 Sheets-Sheet 10 ENGINEAPPARATUS This application is a division of my prior copendingapplication Ser. No. 857,162 filed Sept. 11, 1969 now US. Pat. No.3,614,944, and a continuation-in-part of my copending application Ser.No. 786,233 filed Dec. 23, 1968. This application relates toimprovements to the apparatus shown in prior application Ser. No.786,233. Certain principles and aspects of the present invention arereadily applicabble, moreover, to internal combustion engine fuelsystems other than those shown in my prior application.

It is desirable, after starting, and during normal warmed up operation,that the fuel-to-air ratio of the mixture burned by an engine eitherremain relatively constant or vary slightly in a desired sense over awide range of speed and load conditions. It is generally preferred thatthe ratio either remain constant or vary slightly inversely with enginespeed. Most internal combustion engines are controlled by a primarycontrol which operates basically to control engine torque. In mostcarburetor-equipped engines the engine primary control controls torqueby adjusting a throttle plate to adjust the amount of fuel-air mixtureaspirated into an engine cylinder during a stroke. The amount ofmixtureaspirated per stroke determines the amount of air drawn into thecarburetor air intake per stroke, and because of the venturi principleupon whichthe carburetor operates, the amount of fuel mixed with airvaries rather proportionally. Because carburetor air flow inherentlycontrols carburetor fuel flow, the fuel-to-air ratio of the mixtureinherently tends to remain somewhat constant over a wide range of enginespeeds. Any unwanted variation can be corrected easily in a number ofways, such as by use of a metering needle valve moved in response toengine vacuum.

While fuel injection engine systems have a number of advantages overcarburetor-equipped engine systems for a number of applications, fuelinjection systems do not have the above-described automaticmixturemaintaining tendency of carburetor systems. In fuel injectionsystems engine torque is usually controlled by simultaneous variation ofthe amount of fuel pumped per stroke and the amount of air aspirated perstroke, with the engine torque control linked to both vary the amount offuel pumped per stroke and to vary a throttle plate in the air intakeduct. When a decrease in engine load causes an increase in engine speedat a given torque control setting, the increased air flow causes anincreased pressure drop across the air intake structure, decreasing theamount of air aspirated per stroke, With the amount of air aspirated perstroke decreased and the amount of fuel pumped per stroke remaining thesame, it will be seen that the fuel-to-air ratio disadvantageously tendsto increase as enginespeed increases due to decreased load at a giventorque controlsetting. While such a variation in fuel-air ratio is notregarded as a serious limitation in many applications, in certain otherapplications, and particularly in those where the engine frequentlyoperates under widely-varying load conditions, it is desirable that fueldelivery vary inversely with engine speed, or directly with engine load,as well as with adjustment of the engine torque control, or primarycontrol. The variation of fuel delivery so as to maintain a desiredfuel-to-air relationship may be termed secondary control. it is aprincipal object of the invention to provide improved fuel injectionsystems having effective and reliable secondary control.

Secondary control, or automatic variation of fuel delivery with enginespeed in order to keep the fuel-air mixture relatively constant, isknown'in the prior art in connection with diesel engines and certaingasoline engines which use fuel injection. One prior art system is shownin US. Pat. No. 3,443,554, for example. However, those prior art systemsof which I am aware require a centrifugal governor and/or other verycomplex and expensive mechanisms in order to provide secondary control.A very important object of the invention is to provide secondary controlusing much simpler and more economical apparatus which is easilyconstructed and highly reliable. Use of the present invention allows oneto replace extremely complex and expensive mechanisms with merely aspring-loaded check valve.

In some prior art fuel injection systems themaximum pressures developedin the injection pump during an engine cycle vary appreciably with bothengine speed and the fuel delivery setting of the basic torque control,while the maximum pressure developed in various other systems variessubstantially solely with engine speed and is substantially independentof the fuel delivery setting of the basic torque control. As will beseen below, secondary control of some systems may be obtained solesly byuse of a spring-loaded check valve having a constant spring load, andsecondary control of the other prior systems may be obtained very simplyby use of a spring-loaded check valve together with means for varyingthe spring load as a function of the fuel delivery setting of the basictorque control. Thus further objects of the invention are to provideeffective secondary control for fuel injection systems of both types.

The basic principles of the present invention are readily applicalbeboth to fuel injection systems used with four-cycle engines and thoseused with two-cycle engines, and readily applicable to both fuelinjection systems used with a single cylinder and those used withmulti-cylinder engines, including those having distributor arrangementsto successively feed plural cylinders in a sequence, and further objectsof the invention are to provide apparatus having such versatility.

Several basically-different types of variable-delivery reciprocatingpumps are used in fuel injection systems, and another object of theinvention is to provide secondary control arrangements which are usefulwith each of the different types of pumps. Several different types ofvariable-delivery rotary pumps are also used in fuel injection systems,and a further object of the invention is to provide a secondary controlarrangement which may be used with one or more such pumps. The inventionis applicable to any fuel-injection system in which a pulsating pump isused and/or in which a distributor is used, in such a manner that a peakpressure impulse varying in accordance with engine and pump speed occursduring a pumping cycle.

A further object of the invention is to provide a simple secondarycontrol device which may also be arranged to act as an engine speedgovernor or limiter.

Yet another object of the invention is to provide a single secondarycontrol device which also may be arranged to be responsive to engineacceleration and deceleration to vary the fuel-air ratio in a desiredsense when the engine accelerates or decelerates. A further object ofthe present invention is to provide an improved form of a special typeof rotary pump in which control of delivery is effected by limitedangular adjustment of a substantially stationary member rather than bycontrolling the phase relationship of two rapidlyrotating components.

Other objects of the invention will in part be obvious and will, inpart, appear hereinafter.

The invention accordingly comprises thefeatures of construction,combination of elements, and arrangement of parts, which will beexemplified in the constructions hereinafter set forth, and the scope ofthe invention will be indicated in the claims.

For a fuller understanding of the nature and objects of the inventionreference should be had to the following detailed description taken inconnection with the accompanying drawings, in which:

FIG. 1 is a cross-section view ofa reciprocating variable-delivery pumpof the type shown in my copending application modified to incorporatesecondary control in accordance with thepresent invention.

FIG. 1a shows a modification to the pump of FIG. 1 to provide adifferent type of oil pumping.

FIG. 1b is a porting diagram useful in understanding theoperation of thepump of FIG. 1.

FIGS. 2a and 2b are cross-section views of a rotary variable-deliverydistributing pump of a basic type shown in my prior US. Pat. No.3,057,300 modified to incorporate secondary control in accordance withthe present invention. FIG. 2a is a view taken at lines 2a2a in FIG. 2band FIG. 2b is a view taken at lines 2b2b in FIG. 2a.

FIG. 20 is a view of one plate part of the apparatus of FIGS. 2a and 2b.

FIG. 2d is a view of a portion of FIG. 2a.

FIG. 2e is a view of an adjustable metering plate part of the apparatusof FIGS. 2a and 2b.

FIG. 2f is a timing diagram useful in understanding the operation of theapparatus of FIGS. 2a and 2b.

FIG. 2g is a cross-section view taken at lines 2g2g in FIG. 2d.

FIG. 2h illustrates certain modifications which may be made to thedevice of FIGS. 2a and 2b.

FIGS. 3a, 3b and 3c are diagrammatic views useful in explaining theoperation of the invention with each of three different types ofreciprocating, variabledelivery, constant stroke-length fuel injectionpumps.

FIG. 4 is a cross-section view of a modified form of rotaryvariable-delivery fuel injection pump and distributor.

FIG. 4a is a view of a rotatable distributor plate of the device of FIG.4.

FIGS. 5a to Sc illustrate an alternative embodiment of the invention.FIG. 5a is a cross-section view through the pump, FIG. 5b is a viewtaken at lines Sb-Sb in FIG. 5a, and FIG. 50 is an exploded view inwhich the parts are shown isometrically FIG. 3a diagrammaticallyillustrates the application of the invention to a well-known form offuel injection system using a variable-delivery reciprocating pump inwhich delivery is varied by varying the time or point during the strokeat which an inlet port is closed off. The engine crankshaft ismechanically connected to reciprocate piston P a fixed stroke distance swithin'cylinder CY. Piston P is provided with an axial length whichvaries angularly around the piston, so that rotating piston P to variousangular positions will vary the time or position during each stroke atwhich the piston covers inlet port IP. If port IP is closed off earlyduring the stroke, greater delivery will result. During a rightwardpumping stroke, fuel will be expelled through inlet port IP back to thesupply tank until the piston blocks the port, and forward pumping willoccur through delivery check valve DV to nozzle NO after the pistonblocks the port until the end of the stroke. The angular position ofpiston P is controlled by the engine primary control, which is shown ascomprising an accelerator pedal A. The primary control also varies airintake throttle plate TP via a cam or suitable linkage I to generallyincrease air flow as fuel quantity is increased. At a given setting ofthe primary control, an increase in engine speed due to a decrease inload will decrease the amount of air aspirated per stroke due to therestriction of the air intake structure S, thereby undesirablyincreasing the fuel-air mixture ratio.

As piston P of the injection pump travels rightwardly on a pumpingstroke from the leftward limit position shown in FIG. 3a, fluid willinitially be expelled out inlet port IP. If piston velocity wereuniform, the pressure within the cylinder during the initial travelwould tend to be a constant value dependent upon piston speed and therestriction of unblocked port IP. If piston P is instead reeiprocatedwith simple harmonic motion, or an approximation or modificationthereof, as is usually the case in practical systems using cranks oreccentrics or the like, the linear veloctiy of the piston will insteadvary approximately cosinusoidally, for example, from zero velocity atthe leftward position shown to a maximum speed at mid-stroke, down tozero velocity at the end of the rightward stroke. The linear velocity,assuming simple harmonic motion, may be written as (taking mid-stroke asthe origin):

v=s/2 w COS wt With piston velocity increasing during the initialportion of the stroke, the flow Q through port IP will similarlyincrease in direct proportion. FRom Torricellis theorem Q /A it isevident that the pressure drop across the port, and hence th pressurewithin the cylinder, will vary as the square of the flow Q, and varyinversely with the square of the area of the port, and hence thepressure will increase approximately in accordance with an w (cos cut)characteristic, where w is engine speed in radians per second. Thus thepressure existing in the cylinder when the piston edge reaches the inletport will be higher at higher engine speeds, varying approximately withthe square of engine speed. If the area A of port IP is large, thepressure built up in the cylinder prior to closure of the inlet portwill still be very modest, however, and in many systems is small enoughto be neglected.

If at a given engine speed the piston is rotated to close off port I?later during the stroke, so as to provide lesser delivery per stroke,but still prior to midstroke,

the piston will be seen to reach a greater linear velocity by the timeit begins to close off port IP, and hence a greater pressure will existwhen closure begins than at greater delivery settings. At any givenengine speed, the maximum pressure would be developed if port IP isclosed off approximately at midstroke, when the piston is at its maximumlinear velocity. If piston P is rotated to provide port closure verylate during the stroke, a lesser peak pressure will be developed due tothe lesser piston velocity at the time of closure. Prior art systems ofthe type shown in FIG. 3a did not include the further check valve SV,and its presence should be ignored for the moment. i

As the piston begins to close off port 1?, th pressure furtherincreases,'notonly due to the increase in piston lineanvelocity andincrease in flow, (assuming port closure prior to midstroke) but alsodue to the decrease in unblocked port area irrespective of whether portclosure occurs before or after midstroke. The pressure p rises in thecylinder as the port is closed off roughly in accordance with thefollowing characteristic:

where A, is the initial open area of the port, d is' the axial width ofthe port, and k and k are constants. As the port is gradually closed offit becomes an infinite restriction, and hence as theport is closed offthe pressure rises exponentially, theoretically toward an infinitevalue. However, when the pressure reaches a value determined by thespring loading of delivery check valve DV, the valve DV opens and theincrease in pressure is thereafter limited/Thefrate, at which thepressure rises as the port is closed off depends not only upon thegeometry of the port,-.but also upon the piston speed; The increase inpiston speed not, only increases. the pressure dueto provision ofincreased flow, but al'soincreases the rate of pressure increase bymorequickly closing off the port, so that the rate of pressure-increasevaries as a fairly high power of engine speed. The precise slope of thepressure characteristic will also depend, of' course, upon the shape ofthe inlet port as well as its general width, andthe shape of the pistonedge.

which passes over the port to block the port-As the piston is rotated todecrease the delivery per stroke at a given engine speed, sothat thepistonhas a. greater velocity when it. c losesoff the port, the rate ofpressure increase willbe seen to. increase. Since maximumlinear pistonve-loctiy, occurs at midstroke, the maximum peak pressur'e forlany givenengine speed will occur 'when the portis closed off approximately atmidstroke, when the pump is: adjusted to'pump approximately one-half itsmaximum. delivery perstroke.

The rapid increase in pressure as the portfisiclose'd off applies asudden force to the body of valve' DV, accelerating it rightwardlyagainst the force of the valve spring and providing a damped oscillationof thevalve body. The mass of the valve body, the valve spring, and theviscous resistance of the fuel to motion of the valve bodyafter thevalve is opened will be seen to provide a mass-spring-damper secondorder system. The valve body, eventually returns to a steady-stateposition such that cylinder pressure balances the valve' spring:loading, and cylinder pressure remains substantially at the valuedetermined by the delivery valve spring loading for the remainder of therightward pumping stroke; The

motion of the check valve body required to allow maximum flow throughthe check valve is assumed to be small compared to the length of thecheck valve spring,

and hence the spring may-be assumed to apply a substantially constantforce to the check valve body. Because the pressure drops quickly whenthe delivery check valve opens, the force applied to the valve body hasthe nature of a brief impulse, the amplitude of which varies as anexponential function of engine speed. Thus increased engine speedincreases the amount which the valve body overshoots. After theovershoot, the check valve maintains cylinder pressure substantially ata value determined by the check valve spring loading. The pressure inthe cylinder may increase slightly up to midstroke a's piston velocityincreases and thereafter decrease somewhat as piston.- velocitydecreases during the latter half of the rightward pumping stroke but nofurther sudden increase in pressure will occur during the pumpingstroke. As mentioned above, the foregoing description of operationassumes that check valve SV is not present.

In accordance with the present invention, fuel delivery per stroke maybe descreased with increasing engine speed by provision of the furthersecondary control check valve SV, which is responsive to pump cylinderpressure and operative to spill back increasing amounts of fuel to thesupply as engine speed increases. The ratio between the mass of the bodyof a check valve to the spring force of the spring of the check valvemay be termed the check valve time constant. Secondary control checkvalve SV is provided with a smaller time constant than that of deliverycheck valve DV. As the closure of inlet port I? causes the rapidincrease in cylinder pressure, the pressure is applied simultaneously toboth the delivery check valve and the secondary control check valve. Thepressure temporarily rises above the steady-state delivery valvepressure setting due to the greater inertia or longer time constant ofthe delivery check valve, which delays its opening. During thattemporary high pressure condition the secondary control check. valve SVopens, due to its lesser mass, despite its greater spring loading, .andopening of valve SV spills back fuel to the supplytank and limits thepressure developed in the cylinder. The amount which valve SV opens willbe seen also to depend upon the peak pressure developed in the cylinder,and hence upon engine speed. As well as improving mixture ratio byspilling back some fuel, the quick The variation in volumetricefficiency, or air aspirated per stroke, with speed is ordinarilynon-linear for most engines. Also, the variation of peak pump cylinderpressure with pump speed is non-linear, and the variation in the amountof fuel which a typical springloaded check valve will pass with a givenpressure impulse applied to it is also non-linear. Furthermore, the

peak pump cylinder pressure occuring at a given engine speed varies inaccordance with fuel delivery setting, as described above. Because ofthese varying relationships, it is sometimes difficult to provide adesired fuelair ratio characteristic over widely-varying load conditionsif a fixed spring loading is used on the secondary control check valve.In accordance with a further feature of the invention, the loading onthe secondary control check valve may be varied as a function of theprimary control delivery setting, and in FIG. 3a cam C rotated by theprimary control A is effective to vary the spring load on check valveSV.

If the piston in FIG. 3a is rotated to decrease fuel delivery by closingoff the inlet port later during the first half of the stroke, the pistonwill have a greater velocity as it closes off the port, therebyincreasing the slope of the pressure characteristic, aswill be apparentfrom expression (2), and thereby providing greater impulses to open thesecondary check valve SV. Piston velocity decreases during the latterhalf of the pumping stroke. Thus maximum peak pressure for a givenengine speed is developed if the inlet port is closed approximately atmidstroke, which results when the pump is operating at roughly one-halfof its capacity. Most engine systems require fuel delivery which variesfrom none or some small amount up to a maximum required for normalrunning although even greater delivery may be required for starting.Since minimum delivery requires inlet port closure very late in thestroke, a pump of the type shown in FIG. 3a ordinarily will operate overa range which varies from a minimum delivery condition involving portclosure very near the end of the rightward stroke when piston velocityis low, up to maximum delivery condition involving port closure muchearlier during the stroke when piston velocity is greater. If themaximum fuel required by the engine during running conditions is no morethan half the maximum pump capacity, it will be seen that the peak pumpcylinder pressure developed at a given engine speed will vary directly,though not linearly, with the fuel delivery setting over the entirerunning range of the engine.v Under such conditions, cam C willordinarily provide a spring-loading to valve SV which generallyincreases as the delivery is increased. If the engine requires more fueldelivery than half the pump capacity so that inlet port closure mustoccur prior to midstroke, a plot of the peak pump pressure developed ata given engine speed versus pump delivery will be seen to slopdownwardly at the highest delivery values. Under such an arrangement,cam C will ordinarily provide spring-loading which increases as.delivery is increased upto a given delivery value, after which cam Cwill provide decreasing spring-loading as delivery is further increased.

FIG. 3b diagrammatically illustrates a different for of reciprocatingvariable-delivery constant strokelength injection pump in which deliveryis varied by varying the time during the stroke at which forward pumpingis terminated, rather than varying the time at which it begins. Thepistons P and AP are reciprocated by the engine with some approximationof simple harmonic motion. As piston P travels rightwardly on a pumpingstroke, delivery commences substantially immediately through deliverycheck valve DV, and continues throughout the rightward pumping strokeuntil port TP of auxiliary piston AP registers with port SP of collarCO, at which time fuel is spilled back through hose H to the supplytank. The delivery check valve feeds a nozzle extending into the engineair intake structure in the same manner as in FIG. 3a. The piston Pcontains a bore and a conduit communicating with port T? of auxiliarypiston AP. Collar CO is arranged to be axially adjustable relativetoauxiliary piston AP by means of the engine primary control, so that thetime or position during the stroke at which forward pumping ceases maybe varied to vary the quantity of fuel delivered. Inlet check valve IVadmits fuel to the pump cylinder during the leftward return or suctionstroke.

First consider the operation without the use of secondary control checkvalve SV. At the beginning of a rightward pumping stroke, piston speedbegins at zero and increases cosinusoidally. Pressure builds up in thepump cylinder substantially immediately to a value greater than thesteady-state spring loading of the delivery check valve, and thendecreases to a value commensurate with the delivery check valve loading,as the delivery check valve DV opens. While the velocity of the piston Pis minimum (zero) at the beginning of the stroke, the acceleration ofthe piston is then at its maximum value, and assuming simple harmonicmotion2a=sl2 (1) sin cut. The maximum acceleration of piston P applies amaximum impulse to delivery check valve DV, and the magnitude of theimpulse will be seen to vary as the square of engine speed. As thedelivery check valve opens, the pressure in the cylinder drops markedly.The pressure then increases somewhat until midstroke (assuming collar COis adjusted to provide delivery past midstroke) due to the increasingvelocity of the piston and increased flow through valve DV, but thepressure does not ordinarily approach the initial peak pressure. Whenport TP reaches port SP the pressure drops suddenly and delivery valveDV closes. Inasmuch as the peak pressure occurs at the beginning of thestroke, irrespective of the adjustment of collar CO, it will be seenthat variation of the delivery setting of collar CO by the engineprimary control has no effect on the peak pressure developed within thecylinder.

In accordance with the invention, secondary control check valve SV isprovided in FIG. 3b, again with'a smaller time constant than deliveryvalve DV, so that valve SV opens briefly during the pressure peak tospill back fuel to the supply, and it will be apparent that increasingengine speed causes greater impulses to valve SV, thereby spilling backmore fuel. Because the magnitude of the pressure peaks does not tend tovary with the delivery setting, it is in general less necessary to use acam to vary the spring loading on the valve SV in FIG. 3b. However, theuse of such a cam, in the same manner as in FIG. 3a, allows one to moreeasily tailor the secondary control valve spill-back amount to a givenvolumetric efficiency versus speed characteristic, and

the use of such a cam with the pump of FIG. 3b is within the scope ofthe invention.

While FIG. 3b illustrates a reciprocating variabledelivery pump using aconstant stroke length, its peak pressure characteristic is essentiallythe same as that of a number of reciprocating variable-delivery pumps inwhich the amount of fuel pumped per stroke is varied by varying the pumpstroke length. In such pumps, the peak pressure ordinarily occurs at ornear the beginning of the stroke, and the magnitude of the peak pressuredoes not vary appreciably with the fuel delivery or stroke lengthadjustment. It will be apparent that a secondary control check valve maybe connected to the chamberof such a pump in thesame manner as with thepump of FIG. 3b, with the check valve spring loading being either variedor notvaried as a function of the primary control or stroke lengthsetting.

FIG. 30 diagrammatically illustrates a third form of reciprocatingvariable delivery, constant stroke-length injection pump of a type shownin greater detail in FIG. 1 and also described in detail in my copendingapplication Ser. No. 786,233. Piston P is reciprocated by the enginewith some approximation of simple harmonic motion. A passageway withinpiston P communicates with the pump chamber and selectively communicateswith inlet, port IP and outlet port OP. The passageway edge positionsvary angularly about the piston so that rotation of the piston variesthe time during a given stroke at which inlet port I? is closed 'off andthe time at which outlet port OP is opened, thereby varying the amountof fluid pumped during a rightward pumping stroke. The engine primarycontrol rotatably adjusts piston P to vary pump delivery rate. At theleftward position of the piston inlet port I? is fully opened, and atthe rightward end of the pumping stroke outlet port OP is fully opened.The passageway geometry is arranged relative to the two ports so thatoutlet port OP always opens slightly before inlet port IP is completelyclosed off at any angular position of the piston. With inlet port IPclosing as outlet port OP is opening, the maximum restriction to flowfrom the pump cylinder occurs during the overlap condition when bothports are slightly open.

Consider initially the operation of the pump of FIG. 3c withoutsecondary control check valve SV. As the piston begins a rightwardpumping stroke, fluid is expelled through fully open inlet port 1?, andthe pressure within the pump cylinder remains low. As the inlet portbegins to close off and the maximum flow restriction condition isapproached, the pressure in the pump cylinder increases very rapidly,and then as the maximum restriction overlap condition is passed andoutlet port is opened wider, the pressure decreases. The magnitude ofthe peak pressure developed in the cylinder will be seen to depend uponboth engine speed, which determines the flow rate out of the pumpcylinder, and upon the minimum total open area of the two ports whenboth are slightly open. As the piston is rotated to vary the delivery,it will be seen that the time during the stroke at which the maximumrestriction condition occurs will vary, and if the same maximumrestriction condition occurs at different piston velocities, whichprovide different flow rates from the cylinder, it may be seen that thepeak pressure obtained will also vary with the engine primary controlsetting. If the same maximum restriction condition, i.e., same minimumopen area during overlap, is made to occur for all delivery settings,the peak pressure at a given engine speed will be seen to be obtained ifthe maximum restriction condition occurs substantially at midstroke,when piston linear velocity is greatest, so that the peak pressure for agiven engine speed will occur when the pump is adjusted to pumpapproximately one-half of its maximum delivery per stroke. I

In accordance with the invention, secondary control check valve SV isconnected from the pump cylinder to spill fuel during the occurrence ofthe pressure peaks. One advantage of the pump of FIG. 30 over those ofFIGS. 30! and 3b is that the delivery check valve DV maybe very lightlyloaded, since delivery cannot begin until output port OP is opened,irrespective of pump speed and delivery setting. Furthermore, while thepeak pressure impulse developed in the pumps of FIGS. 3a and 3b andapplied to their secondary control check valves is limited by theopening of their delivery check valves, the peak pressure developed inthe pump of FIG. 30 is substantially independent of the delivery checkvalve loading, and thus the secondary control check valve used in thearrangement of FIG. 3c need not have a shorter time constant than thatof the delivery check valve or otherwise be adjusted relative to anyother check valve.

The above description of FIG. 30 assumes that the same maximumrestriction condition occurs during the overlap condition at all angularadjustments of the piston. By suitably shaping and/or slanting the edgesof the parts relative to the piston passageway edges one can cause thearea of the maximum restriction to vary as the piston is rotated toprovide different delivery rates, and hence one can make the amplitudeof pressure peaks occurring in the pump of FIG. 30 either more afunction of, or less a function of, the delivery setting in whatevermanner one chooses. If the minimum port area occurring during overlap iscaused to increase somewhat with delivery setting up to one-half of pumpcapacity, thereby decreasing the maximum restriction with an increaseddelivery setting up to one-half pump capacity, (and thereafter todecrease with increased delivery if more than one-half pump capacity isused) the magnitude of the pressure impulses will tend to vary less withdelivery setting. If the minimum port area during overlap is caused tovary roughly as a sine-squared function with the delivery setting, itwill be seen that the magnitude of the pressure peaks occurring at agiven engine speed can be made theoretically independent of the deliverysetting, so that no variation in check valve spring loading withdelivery setting is necessary. Because the pump of FIG. 3a requires thatthe inlet port be fully closed, providing an infinite restriction at alldelivery settings, the magnitude of the pressure impulses occurring insuch a pump varies markedly in accordance with the delivery setting,since the delivery setting determines the time during the stroke, andhence the piston velocity at the time the restriction is imposed. Thepump of FIG. 30 (and FIG. 1), by not providing an infinite restriction,but inStead a controllable partial restriction, the minimum area ofwhich can be made to vary with delivery setting, therefore has themarked advantage that the magnitude of the pressure impulses occurringat a given engine speed may be arranged to vary with delivery setting inaccordance with any desired function, or if desired, arranged not tovary appreciably at all. i

It has been shown that while the peak pressure developed during apumping stroke at a given engine speed varies with primary controlsetting with the pump of FIG. 3a, with maximum pressure being developedwhen this pump is pumping at roughly one-half its maximum capacity, thatthe peak pressure developed in the pump of FIG. 3b tends to be largelyindependent of the primary control setting, and that the peak pressuredeveloped at a given engine speed in the pump of FIG. 3c may or may notvary appreciably with delivery setting, depending upon whether its portgeometry is arranged to provide a minimum restriction area which varieswith delivery setting. The amount of fuelspilled back by the secondarycontrol check valve of any of the three systems of FIGS. 3a, 3b and 3cvaries with the peak pressure impulse applied to the check valve in amanner dependent upon the check valve passage geometry, as well as uponits inertia and spring loading.

FIG. 1 illustrates in a cross-section view a form of injector pumpdisclosed and described in detail in my copending application Ser. No.786,233, with certain modifications made thereto in accordance with thepresent invention. The pump is of the basic type described above inconnection with FIG. 3c, .but shown adapted for two-cycle engine use topump oil as well as fuel. The pump comprises a generally-cylindricalcentral casting 120 having a rear head 121 and a front head 122 boltedthereto by means of bolts (not shown), with a suitable gasket (notshown) preferably provided between each head and the central casting.Shaft 123 rotated by the engine crankshaft carries eccentric cam 127.Rotation of cam 127 reciprocates tappet 81, which is carried in bushing82 with an O-ring seal 83a. The right end of tappet 81 bears against theleft end of piston 130, which reciprocates within sleeve 129a. A spring133, only a portion of which is shown, is inserted between head 122 anda right-end face of piston 130 and operates to return piston 130. Alower gear sector 83 pinned to piston 130 is engaged by upper gearsector 84 pinned to control shaft 131, so that rotation of shaft 131angularly positions piston 130. Shaft 131 is angularly positioned byaccelerator pedal or throttle control 103 via arm 104 and a suitablemechanical linkage shown merely as a dashed line. Upper gear sector 84is axially wider than lower gear sector 83 so that the gear sectorsremain enmeshed as sector 83 reciprocates with piston 130.

Oil is supplied from an oil supply-tank (not shown) to chamber 128 via acheck valve (not shown) and a pipe connection made at 128a on the sideof central casting 120. Fuel is supplied to chamber 164 from fuel tank146 via conduit 1450. Oil and fuel inlet ports are provided in sleeve129a at 134 and 136, respectively, and oil and fuel outlet ports areprovided at 135 and 137. Oil piston 161 is urged rightwardly againstfront head 122 by inner coil spring 162. Holes drilled in main casting120 at 142a and 147a connect the outlet ports withlongitudinally-extending passages in which check valves 150 and 151 arelocated, and plugs 1420, 147c close the ends of passages 142 and 147a.Check valves 150 and 151 at the outlet side of the injector pump eachcommunicate with mixing chamber 143 provided in front head 122. TwoV-shaped grooves are milled across the outer periphery of piston 30 asshown by dashed lines at 138 and 139. The bottom of V-groove 138communicates with oil chamber 155 inside piston 130, and the bottom ofV-groove 139 communicates with fuel chamber 160 situated to the right ofpiston 130 and partially within piston 130. At various axial positionsof piston 130 V-groove 138 connects chamber 155 to only chamber 128 viaoil inlet port 134, or to both inlet chamber 128 via inlet port 134 andto mixing chamber 143 via outlet port 135 and check valve 150, or toonly mixing chamber 143 via outlet port 135 and check valve 150. Atcorresponding axial positions of piston 130, V-groove 139 connects fuelchamber 160 to only chamber 164 via inlet port 136, or to both chamber164 via inlet port 136 and mixing chamber 143 via outlet port 137 andcheck valve 151, or to only mixing chamber 143 via outlet port 137 andcheck valve 151. Inlet ports 134 and 136 and outlet ports 135 and 137each comprise an opening which extends partially around sleeve 129a,with each such slot having a uniform dimension measured in the axialdirection of sleeve 1290.

The cutting of V-shaped grooves on the periphery of cylindrical piston130 gives the grooves a width which varies with the angular position ofthe groove around the piston. As is described in greater detail in mycopending application Ser. No. 786,233, varying the angular position ofthe piston within sleeve 129a by means of control shaft 131 varies thetime during a given piston stroke at which the V-grooves willcommunicate with the outlet ports and the time at which the V-grooveswill be cut off from the inlet ports, and hence determines the amount offuel and oil which the pump will pass to the mixing chamber during thepiston stroke.

Piston 130 is shown at its leftmost position in FIG. 1. As piston 130 isurged rightwardly on a pumping stroke, at the beginning of the strokeV-groove 138 connects oil piston chamber 155 via inlet port 134 tochamber 128 so that oil within chamber 155 is expelled from chamber 155back into chamber 128, and V groove 139 connects fuel chamber 160 viainlet port 136 to fuel chamber 164, so that fuel is expelled fromchamber 160 back into chamber 164. At an intermediate time during thestroke determined by the angular position of piston 130, the V-groovesfirst reach and unblock outlet ports and 137 and then move out ofcommunication with and block inlet ports 134 and 136. Provision of suchan overlap condition with the outlet ports always slightly openingbefore the inlet ports are fully closed prevents damage due to fluidblockage. Thereafter during the rightward pumping stroke, as the inletports fully close and the outlet ports increasingly open, oil isexpelled from chamber via outlet port 135, and fuel is expelled fromfuel chamber via outlet port 137, and the fuel and oil mix in mixingchamber 143. The mixing chamber connects to a nozzle (not shown) whichinjects the fuel-oil mixture into the engine air intake duct. Asmentioned in my prior application, the fuel and oil are not mixed in amixing chamber in some applications, and instead, only the fuel is pipedto the nozzle and the oil is pumped to various oil holes at desiredlubrication points within the engine.

The basic pump of FIG. 1 and FIG. 30 differs markedly from many somewhatsimilar prior art fuel metering pumps in that an inlet port is closedand a separate outlet port is opened during a pumping stroke, while theprior art generally (e.g. FIGS. 3a and 3b) has left each pump chamber inconstant communication with an outlet check valve during the entirepumping stroke, so that forward pumpingpast a prior art check valveoccurs either immediately (FIG. 3b) or as soon as the inlet port isclosed off to prevent return pumping (FIG. 3a). If the fluid supply haspositive pressure, the check valve in such prior systems must be loadedto at least the same pressure in order to prevent forward pumping priorto complete closure of the inlet port. And even if the fluid supply isnot pressurized, the pressure in the prior artpump chambers necessarilybuilds up prior to complete closure of their inlet ports, in amountsdependent upon pump speed and dependent upon the amount of restrictionto return flow between the pump chamber and the fluid supply, with theamountof said restriction increasing from a basic amount to completeblockage as the inlet port is gradually closed off. If for ward pumpingis not to occur prior to complete closure of the inlet port, the checkvalve in the prior systems must be loaded to the highest such pressurewhich may occur prior to inlet port closure. The heavier check valveloading necessarily results in higher pressures in the pump chamber,thereby requiring a more precise piston-cylinder fit. In the pump ofFIGS. 1 and 3a, forward pumping cannot occur prior to opening of anoutlet port, irrespective of whether the supply is pressurized, andhence the instant at which forward pumping begins during a pumpingstroke remains substantially independent of pump speed and outlet checkvalve loading, making the quantity of fluid delivered per strokesimilarly independent of pump speed and check valve setting.

In accordance with the embodiment of the present invention illustratedin FIG. 1, fuel chamber 160 is connected to fuel inlet chamber 164 via aspring-loaded check valve 163, the spring loading of which is shown madevariable as a function of control shaft 131 position, by means of cam131a carried on control shaft 131. Rotation of control shaft 131, as bymeans of accelerator pedal 103 and arm 104, so as to rotate piston 130to increase oil and fuel flow rates causes cam 131a to vary the springloading on check valve 163. The precise shape of cam 131a will dependupon the desired variation of fuel-air ratio, the variation in air flowwith engine speed due to the engine air intake structure, the variationof pump cylinder peak pressure with engine speed, the variation of pumpcylinder pressure with delivery setting, and the variation in the amountof fuel spilled back through check valve 163 with peak pressure, all ofwhich determine the variation in the amount of fuel spilled back for agiven engine speed with a given primary control delivery setting. Insome embodiments of the invention, the spring loading of check valve 163need not be varied as a function of throttle position. In thoseembodiments can 131a may be eliminated and check valve 163 held inposition with a fixed spring loading by a plug in head 122. Thepassageway which includes a check valve 163 extends generally in adirection so as to intersect shaft 131 if cam 131a is used. If no cam isused it will be apparent that the passageway may extend out radially inanother direction, such as perpendicularly to the plane of FIG. 1.

In the pump of FIG. 1 the inlet and outlet ports are spaced relative totheir respective V-grooves so that maximum restriction to flow from eachV-groove occurs during the intermediate or overlap interval when eachV-groove slightly communicates with both its inlet port and its outletport. Therefore, the maximum pressure which occurs in pump cylinder 160during a pumping stroke occurs during that intermediate or overlapinterval when both inlet port 136 and outlet vary with pump pistonspeed. The pressure in chamber 160 will be seen to drop from its maximumvalue as piston 130 thereafter continues to travel rightwardly andoutlet port 137 increasingly unblocked.

As was explained above in connection with FIG. 3c, the maximum peakpressure developed in the pump cylinder for any given engine speed tendsto occur if the maximum restriction condition when both inlet an outletports are slightly open occurs when the piston has maximum linearvelocity. Maximum piston velocity usually occurs somewhere nearmidstroke if an approximation of simple harmonic motion is used toreciprocate the piston, and adjustment of the pump to cause the twoports to overlap around the midstroke causes the pump to operate atapproximately one-half its maximum capacity. If one-half or less of thepump maximum capacityis sufficient to supply the maximum fuelrequirements of the engine, the overlap will occur during the last halfof the pumping stroke, and if the shape and spacing of ports 136 and 137and V-groove 139 provide the same minimum area restriction as piston 130is rotated to give different delivery rates, increasing the primarycontrol setting to call for increased delivery will increase the peakpressure developed for a given engine speed and tend to increase theamount of fuel spilled back by the secondary control check valve 163,and in such an arrangement cam 163 may be shaped to provide an increasein check valve spring loading as the primary control setting is adjustedto provide greater fuel flow. If, on the other hand, the maximum fuelrequirements of the engine require more than one-half pump capacity, sothat forward pumping is sometimes required during the first half of thestroke, and the port and V-groove geometry again provides the sameminimum area restriction at different angular positions of piston 130,the cam may be shaped to increase check valve spring loading until theprimary control is adjusted to the midstroke overlap condition, andthereafter to decrease the spring loading as greater amounts of fuel arecalled for. However, if rotation of piston 130 is arranged to vary theminimum area of the maximum restriction which occurs during the overlapcondition, the maximum pressure developed in cylinder 160 can be made tovary directly with delivery setting, or not to vary appreciable withdelivery setting, or even to vary inversely with delivery setting, ifdesired. If the maximum pressure does not vary apprecialby with deliverysetting, it will be apparent that variation of the spring loading oncheck valve 163 becomes unnecessary.

FIG. 1b contains three unrolled or developed views illustrating thegeometry of V-groove 139 relative to ports 136 and 137. Angularadjustment of piston 130 to provide different delivery rates amounts tovertical displacement of the V-groove in FIG. lb relative to ports. Oneach pumping stroke V-groove 139 moves rightwardly relative to the portsfrom a beginning position in which the V-groove is centered on the inletport port 137 are both only slightly open, so that maximum restrictionto flow from chamber is provided, and the magnitude of the maximumpressure will be seento 136. V-groove 139 is shown at I in a minimumdelivery position at the time during the overlap condition when p itleast registers with inlet port 136 and outlet port 137, at II in amedium delivery position at the time during the overlap condition whenit least registers with the ports, and at III in a maximum deliveryposition at the time when it least registers with the ports. It will beseen that the minimum overlap area varies from a small area in I, to alarger area in II, and then to a smaller area at III. Times t,, t and 2indicate the times after the beginning of the pumping strokeat which themaximum restriction occurs under the three different deliveryconditions. Thus it will be seen that the amount of maximum restrictionvaries from a minimum at low delivery rates up to a maximum atapproximately one-half capacity, down to a minimum at maximum delivery.Since piston speed at the time of the overlap condition varies inapproximately the same manner, it will be apparent that the variation inrestriction may be used to offset the variation in piston speed at thetime of overlap, so that the magnitude of the pressure impulsesdeveloped at a given engine speed tends to be largely independent of thepump delivery setting.

If cam 131a is eliminated and a constant spring load is used on checkvalve 163, and if the port geometry provides the same minimumrestriction at different delivery settings, the amount of fuel spilledback through the check valve at a given engine speed will increase withthe primary control delivery setting as the primarycontrol is variedfrom minimum flow to one-half pump capacity, thereby leaning out thefuel-air ratio, and as the primary control delivery setting is furtheradvanced at the same engine speed to provide greater flow than one-halfpump capacity, the amount of fuel spilled back through the check valvewill decrease, thereby providing an increasingly-enriched mixture atincreasing delivery settings.

In the arrangement shown in FIG. 1, wherein increasing fuel spill-backoccurs at increasing engine speeds due to light load conditions but nocomparable oil spillback occurs it will be seen that the amount of oilpumped per stroke remains substantially constant, thereby providinglarger oil-to-fuel and oil-to-air ratios during higher speed-lighterload conditions. Such operation is wholly satisfactory for manytwo-cycle engine applications, and particularly in those two-cycleengine applications where the oil is not mixed with the fuel but insteadpumped to various lubrication joints within the engine.

If desired, the oil-to-fuel and oil-to-air ratios may be tailored byproviding a secondary control oil check valve in similar fashion tospill back oil in amounts varying with engine speed.

While FIG. 1 illustrates a system which dispenses metered amounts of oilas well as fuel, such as is used with two-cycle engine systems, it isimportant to recognize that the invention is in no way restricted tofuel injection systems which dispense two fluids, and is quite asapplicable to four-cycle engine systems wherein oil is not injected intotheengine.

While the mixing and metering pump of FIG. 1 uses separate oil and fuelpistons (161 and 130) to pump oil and fuel with a desired ratio, analternative embodiment shown in FIG. 1a dispenses with the need for aseparate oil piston, and the need for V-groove 138 on piston 130 and theneed for oil inlet and outlet ports 134 and 135 in sleeve 129a. In FIG.1a oil is supplied to oil chamber 128 via an inlet conduit 1280 whichcarries duckbill check valve 601. As cam 127 moves fuel piston 130 on arightward fuel-pumping stroke, thereby increasing the volume of chamber128, oil is drawn into chamber 128 through check valve 601. As spring133 moves piston 130 leftwardly, thereby decreasing the volume ofchamber 128, oil is expelled past oil outlet check valve 150 to mixingchamber 143. The amount of oil which is drawn into chamber 128 during arightward stroke and dispensed to the mixing chamber on the pistonreturn stroke depends upon the cross-sectional area of piston times thelength of the pistonstroke, less the cross-sectional area of tappet 81times the same stroke length, since tappet 81 increasingly enterschamber 128 from bushing 82 as piston 130 increasingly leaves chamber128. If tappet 81 is very slightly less in diameter than piston 130,very little oil will be pumped comparedto the amount of fuel pumped. Itwill be seen that a constant amount of oil will be pumped per stroke,irrespective of the adjustment of control shaft 131. In a variety ofsystems, and in particular those which drive constant loads, it isconsidered unnecessary to maintain a constant fuel-oil mixture ratio.The pump in FIG. 1a is shown without the cam 131a and check valve 163utilized in FIG. 1 to provide secondary control, and such a featureobviously can be added to FIG. la, if desired.

The pressure-responsive secondary control concept ofthe presentinvention is not limited to use with reciprocating fuel injection pumps,and is also applicable to rotary fuel injection pumps of the type whichincorporate distributors to distribute fuel to different enginecylinders. FIGS. 2a and 2b illustrate an application of the invention toa rotary fuel injection pump and distributor device of a basic typeshown in my prior U.S. Pat. No. 3,057,300.

. The pump includes a main casting 201 and a head 202 bolted to casting201 by means of bolts 203,203. Main casting 201 includes a cylindricalbore along axis x-x having three different diameters indicated at 206a,206b and 2060. Drive shaft 207 extends through the bore, beingjournalled in portion 206a of the bore by means of bearings 208a, 208b.Seal retainer washer 210 and seal 211 seal the outer end of shaft 207,and ring 212 carrying seal 213 and rubber O-ring 214 seal shaft 207adjacent bearing 208a. Aligning pin 201a seats in bores in casting 201and head 202, and passes through slots in ring 234 and plate 225,thereby angularly fixing these parts relative to each other. Ring 212 isstationary, and earn 227 is attached to shaft 207 and rotatabletherewith. Plate 232 is capable of limited angular adjustment about axisxx by means of control rod 204, which is reciprocated by adjustments ofthe engine primary control (not shown). Spacer ring 246 having aslightly greater axial thickness than plate 232 surrounds plate 232 andis angularly held by pin 201a. Provision of spacer ring 246 transmitsthe force of disc spring 221 from ring 234 to head 202, so that plate232 is not clamped tightly between head 202 and ring 234 and can beangularly adjusted easily.

A passageway 216a, 216b in casting 201 connects to the fuel supply (notshown), thereby admitting fuel to a ring-shaped chamber 217 formed by anannular groove around the external periphery of ring 212. A plurality ofpassages 218,218 extend inwardly and axially in ring 212 to permit fuelto flow from chamber 217 through holes in ring 220. A dome-shaped springcap or disc spring 221 retaining O-rings 222 and 223 urges plate 220 andring 212 rightwardly in portion 2061) of the cylindrical bore.Stationary plate 225, which is shown in detail in FIG. 20 is mountedagainst tha back of disc spring 221. Plate 225 is provided with anoversize central bore greater than the diameter of shaft 207, providedwith slot 224 to accommodate aligning pin 201a, and provided with sixholes 225a225fspaced in a circle 60 from each other. The

embodiment shown is designed for use with a sixcylinder engine, andother hole arrangements are provided in plate 225 for other types ofengines. Four holes arranged at 90 from each other would be used with afour-cylinder engine, for example.

The oversize central bores in disc spring 221 and fixed plate 225 permitfuel flow from chamber 217 through the holes in ring 212 and plate 220to a ringshaped chamber 226 formed by a cylindrical recess in rotatablecam 227, which is shown in greater detail in FIG. 2d. Cam 227 (FIG. 2d)includes two portions 227a, 227C of slightly different radius, withtransition slopes 227b, 227d between the two portions. The twotransistion slopes are located 180 around the cam from each other. Thecam portion 227a of greater radius may be termed the cam lobe. As cam227 rotates clockwise as viewed in FIG. 2d, slope 227b acts as theleading edge of lobe portion 227a, and slope 227d acts as the trailingedge of thelobe. Inlet passage 228 extends radially within cam 227 fromchamber 226 and opens on the side of the cam at trailing edge 227d, andhence it will be seen that trailing edge 227d is in constantcommunication with the fuel supply. Passage 230 extends inwardly withincam 227 from leading edge 227b to where it intersects outlet passage231, an elongated radially extending slot which also extends axiallythrough the cam. With cam 227 mounted adjacent fixed plate 225, as shownin FIG. 2b, it will be seen that cam outlet slot 231 will successivelyregister with individual ones of the six holes 2250-225 f (FIG. 2c) inplate 225, and will not register with any hole in plate 225 at anintermediate angular position between a pair of holes in plate 225.

As seen in FIGS. 2b and 2d, non-rotatable ring 234 surrounding cam 227contains two circular holes 234a, 234b each opening into its circularcentral bore 234s, and two partially-circular crescent-shaped camfollowers 235, 236 having the same axial length as ring 234 and cam 227seat within holes 234a and 234b, respectively. Compression spring 237carried in a bore in ring 234 urges follower 235 counterclockwise inrecess 234a, and compression spring 238 similarly carried in ring 234urges follower 236 counterclockwise in recess 234b, and hence edges 235aand 236a of the cam followers seat against the periphery of cam 227.

With cam 227 in the position shown in FIG. 2d, it will be seen thatleading edge 227b and outlet slot 231 connect to a relatively largechamber bounded by edge 2350 of follower 235 and edge 236a of follower236, and that trailing edge 227d and inlet passage 228 of the camconnect to a relatively small chamber bounded by edge 236a of follower236 and edge 235a of follower 235. The chamber containing trailing edge227d will be smaller than the chamber containing leading edge 227bbecause of the greater size of cam lobe portion 2270 as compared to camrecess portion 2270. As cam 227 rotates slightly more than 180 from theposition shown, thereby moving the cam lobe 227a to decrease the size ofthe initially larger upper chamber, it will be seen that fuel will beexpelled through passage 230 and slot 231, and as cam lobe 227. movesout'of th initially smaller lower chamber, it will be seen that fuelwill be sucked into that chamber through inlet passage 228.

When cam 227 has rotated slightly more than 180 from the position shownin FIG. 2d, leading edge or rise 227b of the cam lobe will rotatefollower 236 clockwise against the force of compression spring 238, andtrailing edge 227d will release follower 235, allowing its compressionspring 237, to rotate follower 235 counterclockwise. Such movement ofthe cam followers, so that their edges 235b, 236b now seat against thecam, will be seen to re-establish a relatively large chamber in front ofleading edge 227b and a relatively small chamber behind trailing edge227d, so that the next half-revolution of the cam again expels fuel outthrough outlet passages 230 and 231 and again draws in fuel throughinlet passage 228. Thus as drive shaft 207 continuously rotates cam 227,fuel is constantly drawn into passage 228 and expelled through slot 231.The pressure of the fuel in outlet slot 231 will be seen to vary as afunction of pump speed. Also, it will be seen that the flow throughpassage 231 will be substantially constant throughout a completerevolution of the pump cam, except for momentary decreases twice duringeach. revolution when cam followers 235, 236 are rotated. The camfollowers are preferably located at an angular position around ring 234so that the followers rotate at two times when cam slot 231 registerswith a particular two of the six holes in plate 225, and then rotationof the cam followers will occur at a low differential pressurecondition, when both cam inlet passage 228 and cam outlet passage 231are connected to the supply.

As cam 227 rotates through a single revolution, slot 231 will be seen topump fuel back to the supply through successive ones of the six holes inplate 225, and forward pumping out to an engine cylinder nozzle canoccur only when slot 231 lies in between a pair of holes and does notregister with any holes in plate 225. When cam slot 231 does notregister with a hole in plate 225, fluid is pumped out of cam slot 231leftwardly as viewed in FIG. 2b, and as fuel is pumped leftwardly out ofslot 231, the proportion of the time it flows to an injection nozzle tothe time it is returned to the supply, is governed by the adjustment ofadjustable metering plate 232. Adustable metering plate 232, as bestseen in FIG. 2e, contains a central bore 232m to slidingly carry plate232 on drive shaft 207 and allow limited angular rotation of plate 232about axis x-x. Plate 232 is shown in phantom in FIG. 2a in placerelative to shaft 207. Control rod 204 passing through bore 241 incasting 201 and fitted with a seal 242 is provided with a yoke andengaging plate 232 by means of cotter pin 243. As control rod 204 isadjusted by means of the engine primary control, plate 232 is angularlypositioned about axis x x relative to the holes in plate 225. Plate 232is provided with six return pumping holes a through f arranged in acircle at one radius from axis x--x, and six forward pumping holes 3through I arranged in another circle at a different radius, with theholes of the two circles angularly staggered, or out-ofphase' with eachother as shown. Outlet slot 231 of cam 227 is provided with a length sothat it may partially register with holes of bOth circles as the camrotates. The position of outlet slot 231 relative to plate 232 at oneangular'position of drive shaft 207 is shown in dashed lines at 231' inFIG. 2e. The width of cam slot 231 is chosen relative to the size andspacing of the holes in metering plate 232 so that slot 231 alwaysregisters at least slightly with either a return pumping hole of theinner circle or a forward pumping hole of the outer circle, and so thatthere is a slight overlap as the slot passes from a hole in one circleto a hole in the other circle. As cam 227 rotates the slot 231 will be

1. A fluid pump, comprising, in combination: means defining a pumpchamber having return and delivery ports; movable means within saidchamber for pressurizing fluid within said chamber; means for supplyingfluid to said chamber; means for cyclically opening and closing saidports in synchronism with the motion of said movable means to providethree distinct conditions including a first condition in which saidchamber communicates solely with a first said port, a second conditionin which said chamber communicates decreasingly with said first of saidports and increasingly with the other of said ports and a thirdcondition in which said chamber communicates solely with said other ofsaid ports, said ports being arranged so that the total area of saidports communicating with said chamber reaches a minimum during saidsecond condition and the pressure in said chamber reaches a maximumduring said second condition; check valve means connected to releasefluid from said chamber to said means for supplying fluid when saidpressure exceeds a predetermined vaLue; and an output conduit connectingsaid delivery port to a utilization device.
 2. A pump according to claim1 in which said means defining said pump chamber comprises a first platewith a recess having an open end and a second plate mounted adjacentsaid first plate, said second plate containing said return and deliveryports, and in which said means for cyclically opening and closing saidports comprises means for providing a relative rotation between saidfirst and second plates so that said open end of said recess cyclicallycommunicates with said return and delivery ports.
 3. A pump according toclaim 2 in which said check valve means comprises a spring-loaded checkvalve mounted in a passageway within said first plate, said passagewaycommunicating with said recess and said check valve being adapted tomove along a valve-operating axis, said valve-operating axis having aradial component of direction, whereby centrifugal force caused byrotation of said first plate varies the opening pressure of said checkvalve.
 4. A pump according to claim 2 in which said check valve meanscomprises a spring-loaded check valve mounted in a passageway withinsaid first plate, said passageway communicating with said recess andsaid check valve being adapted to move along a valve-operating axis,said valve-operating axis having a tangential component of direction,whereby angular acceleration of said first plate varies the openingpressure of said check valve.
 5. A rotary pump comprising a first memberdefining a cylindrical bore; a rotor disc mounted to rotate within saidcylindrical bore and having a periphery provided with two portions ofdifferent radius; a pair of follower means rotatably carried in saidfirst member and spring-biased against said rotor disc to selectivelydivide the space of said bore surrounding said rotor into chambers whichalternately increase and decrease in volume as said rotor disc isrotated, a first passage in said rotor disc connecting the chamber whichis expanding to a fluid source, and a second passage in said rotor discconnecting the chamber which is contracting to outlet port meansextending between opposite sides of said rotor disc, whereby rotation ofsaid rotor disc creates fluid pressure in said outlet port means of saidrotor disc; a first stationary plate engaging one side of said rotordisc, said plate having at least one opening connected to said fluidsource, said outlet port means of said rotor disc being arranged toperiodically communicate with said opening as said disc is rotated; asecond plate engaging the other side of said rotor disc and mounted forlimited angular adjustment, said second plate having at least one returnopening communicating with said fluid source and at least one deliveryopening communicating with a utilization device, said outlet port meansbeing arranged to periodically communicate with said openings in saidsecond plate as said disc is rotated, whereby pressure created in saidoutlet port means may provide fluid flow through one or the other ofsaid openings of said second plate during time periods when said outletport means does not communicate with said openings in said firststationary plate, and the fluid flow occurring during said time periodsis selectively proportioned between said openings of said second plateby the angular adjustment of said second plate, thereby controlling thedelivery of fluid to said utilization device.
 6. A pump according toclaim 5 wherein said openings in said plates are angularly spaced aroundthe axis of rotation of said rotor disc so that said outlet port meanssuccessively communicates with said openings in the following sequence:first with said opening in said first plate, next with said deliveryopening in said second plate and thirdly with said return opening insaid second plate, whereby delivery of fluid to said utilization devicebegins at a predetermined angular position of said rotor disc andterminates at varying angular positions dependent upon adjusTment ofsaid second plate.
 7. A pump according to claim 5 wherein said openingsin said plates are angularly spaced around the axis of rotation of saidrotor disc so that said outlet port means successively communicates withsaid openings in the following sequence: first with said opening in saidfirst plate, next with said return opening in said second plate andthirdly with said delivery opening in said second plate, wherebydelivery of fluid to said utilization device begins at varying angularpositions of said rotor disc dependent upon adjustment of said secondplate and terminates at a predetermined angular position of said rotordisc.
 8. A pump According to claim 5 wherein the sizes and spacings ofsaid openings in said plates are arranged in relation to said outletport means so that said outlet port means communicates with at least oneof said opeings at all angular positions of said rotor disc.
 9. Rotarydistributor pump apparatus for supplying fluid successively and equallyin controlled amounts to a plurality of output conduits, comprising, incombination: pump means having means defining a bore and a rotor adaptedto rotate in said bore about an axis; a fluid source connected to saidpump means, said rotor including an outlet port radially displaced fromsaid axis to sweep in a circular path as said rotor is rotated, rotationof said rotor creating fluid pressure at said outlet port; stationaryplate means having a plurality of openings spaced apart in a circularpath to successively communicate with said outlet port as said rotor isrotated, each of said openings being connected to said fluid source;adjustable plate means capable of limited angular adjustment and havinga plurality return openings spaced apart in a first circular arrangementto successively communicate with said outlet port as said rotor isrotated and a plurality of delivery openings spaced apart in a secondcircular arrangement to successively communicate with said outlet portas said rotor is rotated, each of said return openings being connectedto said fluid source and each of said delivery openings being connectedto a respective one of said plurality of output conduits.
 10. Apparatusaccording to claim 9 in which said outlet port is arranged tocommunicate with at least one of said openings at all angular positionsof said rotor.
 11. Apparatus according to claim 9 in which said rotor isgenerally cylindrical and provided with two end faces, said outlet portextending between said two end faces, said stationary plate means beingdisposed against one of said end faces and said adjustable means beingdisposed against the other of said end faces.
 12. Apparatus according toclaim 9 having a passage in said rotor connecting said outlet port tosaid fluid source, said passage containing a spring-loaded check valve.13. Apparatus according to claim 12 wherein the portion of said passagecontaining said check valve extends at least partially radially in saidrotor, whereby centrifugal force affects the opening pressure of saidcheck valve.
 14. Apparatus according to claim 12 wherein the portion ofsaid passage containing said check valve extends at least partiallynon-radially in said rotor, whereby angular acceleration of said rotoraffects the opening pressure of said check valve.
 15. Appparatusaccording to claim 9 wherein said means defining said bore includes apair of passages connected between said bore and said fluid source, eachof said passages containing a spring-loaded check valve.
 16. Apparatusaccording to claim 15 having spring means acting between said checkvalves to apply equal spring loading to said check valves.
 17. Apparatusaccording to claim 16 having cam means for varying said spring means tovary the spring loading applied to said check valves equally.
 18. Rotarydistributor pump apparatus for supplying fluid successively and equallyin controlled amounts to a plurality of output conduits, comprising, incombination: pump means having means Defining a bore and a rotor adaptedto rotate in said bore about an axis; a fluid source connected to saidpump means, said rotor including an outlet port radially displaced fromsaid axis to sweep in a circular path as said rotor is rotated, rotationof said rotor creating fluid pressure at said outlet port; stationaryplate means having a plurality of return openings spaced apart in acircular path to successively communicate with said outlet port as saidrotor is rotated, said stationary plate means also having a plurality ofdelivery openings spaced apart in a circular path to successivelycommunicate with said outlet port as said rotor is rotated, saidopenings being phased so that said outlet port alternately communicateswith return openings and delivery openings, said return openings beingconnected to said fluid source; a distributor plate arranged to rotatein fixed phase relationship with said rotor, said distributor platehaving port means radially displaced from said axis to sweep in acircular path as said distributor plate is rotated, said port means ofsaid distributor being phased relative to said outlet port of said rotorto communicate with individual ones of said delivery openings when saidoutlet port communicates with said ones of said delivery openings,respectively; and adjustable plate means capable of limited angularadjustment, said adjustable plate means having a plurality of returnopenings spaced apart in a first circular arrangement to successivelycommunicate with said port means of said distributor plate as saiddistributor plate is rotated and a plurality of delivery openings spacedapart in a second circular arrangement to successively communicate withsaid port means of said distributor plate as said distributor plate isrotated, each of said return openings of said adjustable plate meansbeing connected to said fluid source and each of said delivery openingsof said adjustable plate means being connected to a respective one ofsaid plurality of output conduits.
 19. Apparatus according to claim 18in which said rotor is generally cylindrical and provided with two endfaces,said outlet port means extending between said two end faces, saidstationary plate means comprising a first plate disposed against one ofsaid end faces and containing said return openings of said stationaryplate means and a second plate disposed against the other of said endfaces and containing said delivery openings of said stationary platemeans.
 20. Apparatus according to claim 18 in which said outlet port ofsaid rotor is arranged to communicate with at least one of said openingsof said stationary plate means at all angular positions of said rotor.21. Apparatus according to claim 18 having a passage in said rotorconnecting said outlet port to said fluid source, said passagecontaining a spring-loaded check valve.
 22. Apparatus according to claim18 wherein said means defining said bore includes a pair of followermeans spring-biased against said rotor to selectively divide the spaceof said bore surrounding said rotor into chambers which alternatelyincrease and decrease in volume as said rotor is rotated, said borebeing cylindrical and said rotor having a periphery provided with twoportions of different radius, said portions of different radius beingoperative to control said follower means as said rotor is rotated. 23.Apparatus according to claim 18 wherein said bore is provided with twointernal surface portions of different radius and said rotor carries apair of follower means spring-biased to selectively divide the space ofsaid bore surrounding said rotor into chambers which alternatelyincrease and decrease in volume as said rotor is rotated, said internalsurface portions of different radius being operative to control saidfollower means as said rotor is rotated.
 24. Apparatus according toclaim 18 Wherein said openings in said plate means are angularly spacedaround said axis so as to provide fluid flow from said outlet port ofsaid roTor in the following sequence: (1) through one of said returnopenings of said stationary plate means, (2) through one of saiddelivery openings of said stationary plate means, through said portmeans of said distributor plate, and through one of said deliveryopenings of said adjustable plate means, and (3) through said one ofsaid delivery openings of said stationary plate means, through said portmeans of said distributor plate, and through one of said return openingsof said adjustable plate means, whereby delivery of fluid to the outputconduit connected to said one of said delivery openings of saidadjustable plate means begins at a predetermined angular position ofsaid rotor and terminates at varying angular positions dependent uponadjustment of said adjustable plate means.
 25. Apparatus according toclaim 18 wherein said openings in said plate means are angularly spacedaround said axis so as to provide fluid flow from said outlet port ofsaid rotor in the following sequence: (1) through one of said returnopenings of said stationary plate means, (2) through one of saiddelivery openings of said stationary plate means, through said portmeans of said distributor plate, and through one of said return openingsof said adjustable plate means, and (3) through said one of saiddelivery openings of said stationary plate means, through said portmeans of said distributor plate, and through one of said deliveryopenings of said adjustable plate means, whereby delivery of fluid tothe output conduit connected to said one of said delivery openings ofsaid adjustable plate means begins at varying angular positions of saidrotor dependent upon adjustment of said adjustable plate means andterminates at a predetermined angular position of said rotor. 26.Apparatus according to claim 18 having a passage in said distributorplate connecting said port means of said distributor plate to said fluidsource, said passage containing a spring-loaded check valve. 27.Apparatus according to claim 26 wherein the portion of said passagecontaining said check valve extends at least partially radially in saidrotor, whereby centrifugal force affects the opening pressure of saidcheck valve.
 28. Apparatus according to claim 26 wherein the portion ofsaid passage containing said check valve extends at least partiallynon-radially in said rotor, whereby angular acceleration of said rotoraffects the opening pressure of said check valve.